Generally, this invention relates to fluid handling methods and apparatus usable to enhance the performance of centrifugal fan systems. Specifically, the invention focuses on fluid handling methods and apparatus that involve a novel fluid diffuser that can be used to increase the static pressure of an impelled fluid beyond that increase observed using conventional diffusion methods and apparatus. A preferred embodiment involves a vaneless diffuser that converges air passing through it as it radially extends an interface through which this air is output to a downflow air handling environment.
As a brief technical overview, a centrifugal fan discharge has both a radial (e.g., in a direction perpendicular to the axis of rotation of the impeller) and usually also a tangential velocity component (e.g., tangential to a curve such as a circle traced by the rotating impeller); an axial fan discharge has both an axial (e.g., parallel with the axis of rotation of the impeller) and tangential velocity component; a mixed flow fan discharge has tangential, radial and axial velocity components.
Centrifugal fans exist in a variety of configurations. They may be either contained or housed within scrolls (e.g., circular scrolls) for pressure recovery or direct connection to a duct system, or un-housed (e.g., un-scrolled) for use in pressurizing plenums or large volumes. Pressure recovery in a scroll generally refers to recovery of static pressure upon a decrease of an air speed in a direction parallel with a centerline of the flow area of the scroll's substantially uni-directional diffusing section. Centrifugal fans may be further distinguished among themselves by the discharge angle of the fan blades relative to the radial direction. Radial blades discharge fluid (including gas, which itself includes air) in the radial direction. Backward curved blades cause fluid to discharge more in a direction opposite rotation and produce the highest static pressure (for a given amount of input work) as compared with other blade configurations. Forward curved blades discharge fluid more in the direction of fan rotation and have the highest tangential discharge velocity and the smallest static pressure production for a given rotational speed and fan diameter. In other words, as compared with backward curved fans, more of the work done on the fluid by forward curved fans is observed as fluid velocity instead of static pressure.
The desire to maximize the increase in static or pumping pressure of a fluid impelled by a centrifugal fan has been known for many years. It is well acknowledged that it is static pressure and not dynamic pressure of a fluid output by a centrifugal fan system that is more valuable and useful for the intended purposes of most if not all centrifugal fan applications (e.g., supplying air to ducts for eventual release to rooms in a building). Conventional attempts to increase the amount of fluid energy observable as static pressure have resulted in scroll diffusers that seek to increase the cross-sectional flow area of the scroll's unidirectional diffusing section so as to cause a decrease in the speed of the fluid that is parallel with the centerline of the flow area of the scroll's diffusing section, and thereby decrease the dynamic pressure of the fluid. This decrease in dynamic or velocity pressure results in an increase in the static pressure of the fluid because of conservation of energy principles, (see, e.g., U.S. Pat. No. 6,185,954). However, such a diffuser is not without its problems. Not only is it limited in application to scrolled fans and ducted collection systems, but it typically requires a diffuser (e.g. a “jetting” extension) that is so long (e.g., several times the diameter of the fan) that it complicates installation. Maldistribution of airflow often observed in the ducted diffuser section may also lead to less efficient conversion of velocity to static pressure.
Unhoused centrifugal fans, called plenum fans or, when a backwardly inclined airfoil blade is used, plug fans, are also used in many applications for ventilation and material handling (e.g., the pumping of solid materials such as sawdust). These fans are installed in relatively large volumes such as plenums that may be several times the diameter of the fan. It is important to take note of the prevailing attitudes towards opportunities to recover static pressure in unhoused centrifugal fans. As reported in literature describing the application of such centrifugal fans (ASHRAE Journal, October 1997, C. W. Coward; Pace Company Technical Report, April 1995) the velocity pressure produced at the discharge of these unhoused (e.g., unscrolled) “plug” fans is “for all practical purposes, zero” (due to the large outlet area of such unscrolled fans), and therefore is not available for transformation or conversion so as to increase the static pressure of the discharge. The Pace Company document further states that:                There is no static pressure regain when using an un-housed plug fan and Pv equals zero . . . Documents which indicate efficiencies near or above 80% are most certainly based on tests of a fan wheel in a scroll. In order to achieve the submitted efficiency, a scroll must be employed . . . . Pace's opinion is that the absence of a scroll housing limits the mechanical efficiency of a plug fan to somewhere in the low 70's. It is quite doubtful that one exists which performs much better. It is, therefore, our recommendation that uses of un-licensed products with efficiencies in excess of 75% should be avoided unless some clearly identifiable innovation or design change has been implemented. ASHRAE Journal, October 1997, C. W. Coward; Pace Company Technical Report, April 1995These comments reflect the prevailing opinion of those experienced in the art of centrifugal fans and clearly “teach away” from the invention described herein by suggesting that increasing the static pressure of an unscrolled centrifugal fan by capturing energy from the fan's outlet velocity is simply not possible. However, at least one embodiment of the present invention increases the static pressure of an unscrolled centrifugal fan by doing precisely that—converting the fan's outlet velocity energy (specifically the tangential velocity pressure) to static pressure. This manner of fan performance improvement is in direct contravention to the prevailing opinion of those experienced in the art, as expressed in the Pace Company document. In general it appears that there are no devices currently in use or discussed in the literature that allow recovery of velocity pressure from plenum or plug fans to the degree now possible. At least one embodiment of the present invention effects an efficiency in excess of 75%, and indeed in excess of 80%, without the use of a scroll and its related disadvantages. Indeed, “a clearly identifiable innovation or design change” inheres in the instant inventive technology.        
Vaneless diffusers have been the subject of analysis and experimentation as applied to centrifugal compressors since the 1940's (see, e.g., J. D. Stanitz, NACA TN 2610 (1952); J. P. Johnston, “Losses in Vaneless Diffusers of Centrifugal Compressors and Pumps” ASME Journal of Engineering for Power, 1966; H. S. Dou, “Analysis of the Flow in Vaneless Diffusers with Large Width-to-Radius Ratios”, ASME Journal of Turbomachinery, 1998). However, none of these references discloses or investigates the optimization of vaneless diffusers to effectively recover velocity pressure. The central focus appears merely to be the unsteady flow behavior in vaneless diffusers at the onset of rotating stall Where vaneless diffusers are mentioned in conjunction with a centrifugal fan, there is no disclosure relative to optimization of vaneless diffusers (via, e.g., axially converging oppositely facing diffuser forms as a radial distance from a centrifugal fan rotation axis increases) to effectively recover static pressure from centrifugal fans.
The mechanics of vaneless diffusers applied to centrifugal compressors or pumps has been documented in the technical literature (see, e.g. Diffuser Design Technology, David Japikse, 1998 and other papers referenced above). As pointed out above, conventional thinking (as indicated by the 1995 Pace Co. technical report) was that fluids discharged by unscrolled centrifugal fans did not present an opportunity to increase static pressure. As such, vaneless diffusers used with centrifugal fans were designed merely to prevent rotating stall, e.g., and were not in any way shaped to optimize and/or enhance velocity pressure recover. Indeed, the only known vaneless diffusers used with centrifugal fans are parallel plates. (see, Tsurusaki, H., et al., “A Study on the Rotating Stall in Vaneless Diffusers of Centrifugal Fans”, ISME International Journal, 1987, Vol. 30, No. 260., pp. 279–287).
But as reported in the literature (See Japikse, supra, or NACA TN2610, 1952), such vaneless diffusers are relatively inefficient when applied to pumps or compressors (each of which have significantly higher operative pressure regimes than those of centrifugal fans and are designed to operate on primarily radial flow). Essentially, boundary layer effects dominate the flow field inside the centrifugal compressor's diffuser and lead to flow separation and reversal, and higher viscous losses because of the relatively narrow flow path
Examples of vaneless diffusers applied to centrifugal compressors include U.S. Pat. No. 6,382,912 (Lee and Bein), which disclosed a particular wall contour having a pinchpoint for optimizing the performance of a vaneless diffuser connected to a compressor. U.S. Pat. No. 6,382,912 relies on the reduction of radial velocity to achieve an increase in static pressure.
A recent analysis (Yu-Tai Lee, “Direct Method for Optimization of a Centrifugal Compressor Vaneless Diffuser” ASME Journal of Turbomachinery, 2001) reported a method for optimizing a vaneless diffuser for centrifugal compressors. The technique reported is embodied in the above-mentioned U.S. Pat. No. 6,382,912. The flow regime imposed by the compressor is predominantly radial and compressible (Mach number in excess of 1.0), and the flow passage is extremely narrow compared to the diffuser's length of radial extension. The optimization approach is based on fixing the outlet dimensions of the diffuser and optimizing the radial velocity diffusion by optimally shaping one surface of the diffuser. As will be seen by subsequent discussion, this approach is vastly different from that of at least one embodiment of the present invention, in which optimal performance is achieved by adjusting the diffuser contour and outlet dimensions to prevent problems associated with (or related to) recirculation of the radial velocity component while maximally diffusing the tangential velocity component, these problems including but not limited to energy losses. Further, as explained above and below, the centrifugal compressor and the centrifugal fan flow regimes are vastly different.
Centrifugal fans differ from centrifugal compressors in several important ways. First of all, the axial dimension (parallel to the fan's axis of rotation) or axial length of the fan output space (roughly the width of the fan wheel) is significantly larger. As a result, the diffuser flowpath 5 of a centrifugal fan is less dominated by boundary layer effects than in a diffuser used with a centrifugal compressor. Additionally, centrifugal fans operate at speeds and pressures at which the behavior of impelled, flowing fluid 6 (e.g., air 7) may usually be appropriately modeled by ignoring compressibility effects (i.e., assuming an incompressible fluid), in addition to ignoring heat transfer effects. However, such an assumption is entirely inappropriate for centrifugal pumps and compressors.
Other distinctions relative to a centrifugal fan's operative regime as compared with the operative regime of centrifugal compressors are as follows: as but one initial distinction, the typical rotational speed of a centrifugal compressor is orders of magnitude (e.g., 10 times, 100 times) greater than that of a centrifugal fan (a typical upper speed limit of centrifugal fans may be 2000–3000 RPM (revolutions per minute) while a typical speed range of centrifugal compressors may be 10,000 to 100,000 RPM). Centrifugal fans typically effect a static pressure rise (in inches of water) of less than 10 inches, while centrifugal compressors typically effect a pressure rise of greater than 60 inches. Such pressure-related differences constitute one reason why flow behavior of a fluid impelled by a centrifugal fan can often be adequately predicted and/or modeled under an incompressible flow assumption, while such an assumption may be entirely inappropriate in predicting the operative response of a centrifugal compressor, particularly where the fluid is a gas such as air. Compressibility effects become significant (i.e. greater than a 5% change in fluid density for air) at Mach numbers greater than 0.3. According to Japikse (See Centrifugal Compressor Design and Performance, Japiske, 1996) compressors have tip Mach numbers from 0.6 to above 1.0; centrifugal fans have Mach numbers less than 0.3. Such reflects a fundamental difference in the two types of turbomachinery and in the operative response of a fluid impelled by each of them. Further, as indicated, the flow regime in a centrifugal compressor and a centrifugal pump (and through conventional diffusers that may be used in conjunction with them) is primarily radial whereas, in a preferred embodiment of the instant application, the flow regime of and the output from the centrifugal fan (and through at least one embodiment of the inventive diffuser that may be used in conjunction with a centrifugal fan) is primarily tangential.
FIG. 4 shows a Cordier Plot relating dimensionless values of turbomachinery for “well designed” units. It is but one indicator of the fundamental differences of centrifugal fans from centrifugal compressors. The Cordier Plot shows a graph of specific diameter (y-axis) vs. specific speed (x-axis), where specific diameter=(head rise coefficient^¼)/(flow coefficient^½), where head rise coefficient=g(head rise)/(((rpm speed)^2) (diameter^2)), and flow coefficient=flow volume/((rpm speed)(diameter^3)). In the 1950's, Cordier found that well-designed turbo-machines fall on the curve of the Cordier Plot of FIG. 4. As shown by this graph, centrifugal fans typically have a specific diameter of between 2 and 4. Additionally, they have lower pressures than centrifugal compressors, and significantly wider flow paths than compressors and thus, their operation can be improved by the incorporation of a vaneless diffuser. On the other hand, centrifugal compressors, which have specific diameters greater than four and fairly narrow flow paths, are not particularly suited for use with vaneless diffusers. The narrow flowpath of any vaneless diffuser that would be used with the centrifugal compressors would cause significant viscous and frictional losses, thereby compromising any increase in static pressure.
Such above-mentioned fundamental differences alone and in combination render the two types of turbomachinery and the flow behavior of fluid impelled by them sufficiently and fundamentally different, enough so that one would not expect that performance enhancing design features of one of the types of turbomachinery would necessarily enhance performance of the other. Indeed, such would be entirely unexpected.
U.S. Pat. No. 4,323,330 (1982) discloses use of a vaneless diffuser with a mixed flow fan in which impelled air has a radial , axial and tangential velocity. However, the diffuser described in U.S. Pat. No. 4,323,330 relies on changes in effective flow area to reduce axial and radial velocity of impelled air—it does not cause the greater part of its increase in static pressure by reducing tangential velocity—a key feature of at least one embodiment of the present invention. As but a few additional distinctions, the mixed flow fan diffuser of U.S. Pat. No. 4,323,330 does not rely on conservation of angular momentum principles to effect an increase in static pressure (as does at least one embodiment of the present invention); the mixed flow fan diffuser of U.S. Pat. No. 4,323,330 axially diverges air flow (instead of axially converging it as in at least one embodiment of the present invention); the mixed flow fan diffuser of U.S. Pat. No. 4,323,330 includes a partial flow obstructing structure (see parts 48, the “vertically extending orifice portion” and 48c); the mixed flow fan diffuser of U.S. Pat. No. 4,323,330 does not smoothly direct impelled air flow; and the mixed flow fan diffuser of U.S. Pat. No. 4,323,330 generates a flow regime (a mixed flow) that includes an axial component and that is therefore entirely different from the centrifugal fan flow regime of at least one embodiment of the present invention. Even though the mixed fan of U.S. Pat. No. 4,323,330 produces tangential velocity, that patent does not disclose decreasing the tangential velocity to increase static pressure. Instead, its mode of pressure recovery is disclosed by its Diagram b and the related discussion of column 2, lines 10–27, in which there is only reference to the principle of conservation of energy and none to the principle of conservation of angular momentum. That U.S. Pat. No. 4,323,330 does not disclose decreasing the tangential velocity to increase static pressure is particularly evident upon consideration of the patent's disclosure relative to rotating diffuser plates, as such rotating plates would expectedly increase the tangential velocity (in stark contrast to the regain of static pressure effected by a decrease in tangential velocity as seen in the stationary diffuser of a preferred embodiment of the instant invention). Not only does the invention described in U.S. Pat. No. 4,323,330 focus on increasing flow area to recover static pressure from other than tangential velocity, but it does not appear to have the radial extension necessary to reduce tangential velocity, and it does not address controlling the radial velocity in an manner. Indeed, U.S. Pat. No. 4,323,330 illustrates how the manipulation of tangential velocity to increase static pressure was not well considered prior to the present invention.
A clearly evident problem with conventional diffusers may be that none seeks to manipulate both radial velocity and tangential velocity of an impelled fluid output by the centrifugal fan in order to maximize the static pressure recovery, as is seen in at least one embodiment of the instant inventive technology. As such, conventional centrifugal diffusers do not achieve optimal or maximal static pressure recovery.
Vaned diffusers have been proposed for recovery of velocity pressure but have poor off-design performance and as they recover relatively little static pressure, have very low recovery efficiency (which may be defined as the percentage of dynamic pressure at the diffuser inlet that is converted to static pressure). Vaned diffusers are offered commercially in conjunction with centrifugal fans but because of the poor performance discussed above, have not been widely applied.
A common current practice to recover velocity pressure in centrifugal fans is to use curved impeller blades to direct the outlet flow from these fan blades towards a direction opposite fan rotation. This redirection has the effect of reducing the discharge tangential velocity of air leaving the fan and thereby increasing the static pressure produced by the fan. Such fans, called backward inclined or backward curved, produce higher static pressure as compared with that static pressure resulting from fans with blades that are configured in a manner other than backward curved but, because of geometric and practical limitations, still typically produce substantial tangential velocity (regardless of what the Pace Company document states) whose energy is not transformed to static pressure. Relatedly, a disadvantage of this approach is that, in comparison with the approach of at least one embodiment of the instant inventive technology disclosed herein, it requires larger or higher speed wheels to achieve a given static pressure (because as is well understood, the change in total fluid pressure across the fan is proportional to the change in tangential velocity across the fan.).
At least one embodiment of the inventive technology described herein may be applied in any type of centrifugal fan to recover velocity pressure at an enhanced recovery efficiency. However, fans with greater tangential velocities at the discharge (e.g. radial or forward curved fans) offer greater potential for recovery of velocity energy. In addition, the diffuser of at least one embodiment of the present invention can involve shaping, customization or matching to relative to fan characteristics of blade angle, wheel width, and rotational speed in order to perhaps even further optimize the increase in static pressure.